Earth working machine having a shiftable transmission between a drive motor and a rotatable working apparatus

ABSTRACT

An earth working machine ( 10 ), encompassing a drive motor ( 42 ) and a working apparatus ( 22 ) drivable by the drive motor ( 42 ) so as to move rotationally, the drive motor ( 42 ) being connected to the working apparatus ( 22 ), for transfer of a torque, with interposition of a shiftable transmission ( 50; 150 ) comprising at least two gearing stages having different torque transfer ratios, 
     is characterized in that by means of a first gearing stage of the shiftable transmission ( 50 ), an input shaft ( 48 ), coupled on the input side to the drive motor ( 42 ), of the shiftable transmission ( 50 ) is connectable in torque-transferring fashion directly to an output shaft ( 52 ), coupled on the output side to the working apparatus ( 22 ), of the shiftable transmission ( 50 ) so as to rotate together at the same rotation speed; and by means of a second gearing stage of the shiftable transmission ( 50 ), the input shaft ( 48 ) is connectable in torque-transferring fashion, with interposition of a transmission assemblage ( 64 ), to the output shaft ( 52 ) so as to rotate together at different rotation speeds.

BACKGROUND OF THE INVENTION 1. Field of the Invention

The present invention relates to an earth working machine encompassing adrive motor and a working apparatus drivable by the drive motor so as tomove rotationally, the drive motor being connected to the workingapparatus, for transfer of a torque, with interposition of a shiftabletransmission comprising at least two gearing stages having differenttorque transfer ratios.

2. Description of the Prior Art

An earth working machine of the species is known, for example, from DE10 2012 009 310 A1 (U.S. Pat. No. 9,297,125). This document discloses anearth working machine, for example a road miller, recycler, orstabilizer, whose drive train encompasses at one end a diesel internalcombustion engine constituting a drive motor, and at the other end amilling drum, rotatable around a milling drum axis, constituting aworking apparatus. A shiftable transmission that comprises two gearingstages is arranged in the torque transfer path between the internalcombustion engine and the milling drum. Each gearing stage respectivelyencompasses a first gear pair on the input shaft and a second gear pairon the output shaft, the first gear pairs of the two gearing stagessharing, on the input side, a common input drive pinion that is coupledto the input drive shaft of the internal combustion engine so as torotate together at the same rotation speed.

Each gearing stage possesses a dedicated intermediate shaft. Arranged oneach intermediate shaft is a respective output drive pinion, meshingwith the common input drive pinion, of the first gear pair of thegearing stage associated with the intermediate shaft; and a respectiveinput drive pinion of the second gear pair of the associated gearingstage. The input drive pinions of the second gear pair of the first andthe second gearing stage respectively mesh with a separate output drivepinion on an output shaft of the transmission.

Each of the two intermediate shafts is embodied as a split intermediateshaft whose input-side and output-side shaft parts are respectivelydisconnectable, or connectable so as to rotate together, via a shiftableclutch, so that the torque transfer path can be shifted between the twogearing stages under load.

Because each intermediate shaft connects the input shaft of thetransmission to the output shaft of the transmission, at any givenmoment during operation only one of the two intermediate clutches can bein a torque-transferring state, while the respective other intermediateclutch must interrupt the torque transfer path along its intermediateshaft.

The transmission known from DE 10 2012 009 310 A1 (U.S. Pat. No.9,297,125) furthermore comprises a power takeoff to which a hydraulicpump is connected so that it too can be driven by the internalcombustion engine via the transmission.

A further earth working machine is known from DE 10 2012 012 738 A1.

In contrast to the document of the existing art recited earlier, in theearth working machine of DE 10 2012 012 738 A1 torque transfer in theshiftable transmission takes place not via an intermediate shaft butinstead exclusively directly from an input shaft of the transmission toan output shaft, parallel to the input shaft, of the transmission.

Each gearing stage therefore encompasses only one respective gear pair,the input shaft of the transmission carrying both input drive pinions ofthe gear pair, and the output drive shaft carrying the output drivepinions of the two gear pairs. The output drive pinions on the outputdrive shaft are mutually couplable and decouplable fortorque-transferring connection, by way of a claw clutch, exclusively tothe output drive shaft. Here as well, only one gear pair can be coupledin torque-transferring fashion to the output drive shaft at any point intime during operation of the known shiftable transmission.

The input shaft of the transmission passes through the transmissionhousing and, at its longitudinal end pointing away from the internalcombustion engine, projects out of the transmission housing as a powertakeoff. A hydraulic pump to be driven via the power takeoff isrespectively connected at this power takeoff and at a further powertakeoff.

The output drive shaft also projects, at its longitudinal end oppositefrom the coupling to the milling drum, out of the pump housing and formsa combined auxiliary drive/power takeoff. This auxiliary drive/powertakeoff is coupled to a hydraulic pump operable as a hydraulic motor, sothat when one of the two gearing stages is in operation, the outputdrive shaft is rotationally driven by the internal combustion engine,and the internal combustion engine can thus also drive the combinedpump/hydraulic motor apparatus as a hydraulic pump.

The claw clutch used for respective torque-transferring coupling of oneof the two gearing stages can also be brought into a neutral position inwhich neither of the two output drive pinions is coupled intorque-transferring fashion to the output drive shaft. In this neutralposition, the combined hydraulic pump/hydraulic motor apparatus,constituting a hydraulic motor, can rotationally drive the output driveshaft and thus the milling drum for maintenance purposes.

Due to the use of the claw clutch, the shiftable transmission known fromDE 10 2012 012 738 A1 cannot be shifted under load. Instead, in thiscase the load must be briefly interrupted for a shifting operation.

FIG. 12 of U.S. Pat. No. 5,190,398 A discloses an earth working machine,having an internal combustion engine as a drive motor and a rotationallydrivable milling drum, in whose drive train is a shiftable transmissionthat is not characterized in further detail.

The shiftable transmission is coupled for torque transfer on both theinput side and the output side, via respective belt drives, to theinternal combustion engine and the milling drum.

US 2008/0173740 A1 discloses an earth working machine having an internalcombustion engine from whose drive shaft, via a distributortransmission, both a hydraulic pump and (via a separate output drive) amilling drum are rotationally driven. According to the disclosure of US2008/0173740 A1, a stepless transmission referred to as an “infinitelyvariable transmission” (IVT) is located between the internal combustionengine and the milling drum so that the transfer of rotation speed andtorque from the internal combustion engine to the milling drum can bemodified. A steplessly modifiable transmission known as a “continuouslyvariable transmission” (CVT), or an automatically shiftable multi-ratiotransmission, can also be provided instead of the IVT transmission.

The earth working machines of the species having different gear pairsare disadvantageous in that the gear pairs of the shiftabletransmission, which usually step the power supplied from the internalcombustion engine down in terms of rotation speed and up in terms oftorque, exhibit different degrees of wear depending on the operatingtime of the individual gearing stages. The result is either an increasednumber of maintenance operations so that on one occasion gears of theone gearing stage, and on another occasion gears of the other gearingstage, can be maintained or replaced; or that in one maintenanceoperation the gears of all the gearing stages are always replaced eventhough some of them could easily continue to operate given their wearstatus.

SUMMARY OF THE INVENTION

The object of the present Application is therefore to refine the earthworking machine recited previously in such a way that it becomespossible to operate the drive motor with minimal resource utilizationand therefore in a steady state in an optimum rotation speed range witha maximally constant rotation speed, and nevertheless to allow its drivepower to be transferred to the working apparatus, in differentcombinations of rotation speed and torque, without thereby loadingportions of the drive train to different degrees and thus causingdiffering wear.

This object is achieved according to the present invention by way of anearth working machine of the kind recited previously in which, by meansof a first gearing stage of the shiftable transmission, an input shaft,coupled on the input side to the drive motor, of the shiftabletransmission is connectable in torque-transferring fashion directly toan output shaft, coupled on the output side to the working apparatus, ofthe shiftable transmission so as to rotate together at the same rotationspeed; and in that by means of a second gearing stage of the shiftabletransmission, the input shaft is connectable in torque-transferringfashion, with interposition of a transmission assemblage, to the outputshaft so as to rotate together at different rotation speeds.

Thanks to the particular design of the earth working machine accordingto the present invention, either, in a first gearing stage, the inputdrive shaft of the drive motor is fixedly coupled to the output shaft,coupled to the working apparatus, of the shiftable transmission so as torotate together at the same rotation speed (this is referred to in thepresent Application as “direct drive”) or, in the second gearing stage,the input drive shaft and the power of the drive motor available thereare coupled, with a change in rotation speed and torque by thetransmission assemblage, to the output shaft, coupled to the workingapparatus, of the shiftable transmission.

Thus, in the context of the two gearing stages considered here, eitherall the components of the transmission assemblage which are involved ina rotation speed-modifying and torque-modifying power transfer are inthe power-transferring path from the drive motor to the drive apparatus,or none of them are. All the components of the transmission assemblageinvolved in a change in rotation speed and torque therefore wearuniformly, since according to the present invention a shift is made notbetween two transmission sets that transfer torque differently, butinstead between a direct drive mode that dispenses with a transmissionassemblage, and a rotation speed-modifying and torque-modifyingtransmission assemblage.

Because of the particular design of the transmission according to thepresent invention, the latter furthermore has enhanced fail-safecharacteristics as compared with the existing art. This is because evenif the transmission assemblage were to sustain damage, the workingapparatus can continue to be operated by means of the first gearingstage, which directly connects the input and output shafts. The rotationspeed of the working apparatus might then not be optimal for anoperational task to be performed, but in many cases earth working usinga working apparatus driven at a sub-optimal working speed is stillpreferable to total failure of the earth working machine.

In addition to better fail-safe characteristics, a further advantageresulting from the structure of the shiftable transmission of the earthworking machine according to the present invention is enhancedefficiency. Load-dependent losses are unavoidable even with transmissionmeans that transfer torque positively, such as gears. With a tooth set,load-dependent losses constitute a considerable portion (betweenapproximately 85% and 99%) of the total power loss. The first gearingstage, with direct coupling of the input shaft and output shaft of theshiftable transmission, can therefore be operated with considerablydecreased load-dependent losses compared with gear-type transmissionstages of the existing art, since they dispense with any torque step-upor step-down.

As will be explained below, there is no intention here to rule out thepossibility for the shiftable transmission of the earth working machineaccording to the present invention to comprise further power takeoffs inaddition to the output drives recited. The input shaft, coupled to thedrive motor, of the shiftable transmission will therefore be referred tohereinafter as the “principal input drive” of the shiftabletransmission. The output drive, coupled to the working apparatus, of theshiftable transmission is likewise referred to as the “principal outputdrive.”

Although, purely theoretically, the drive motor can be a drive motorbased on any physical operating principles, the advantages of thepresent invention are usable to a particular extent when an internalcombustion engine, in particular a diesel internal combustion engine, isused as a drive motor, since the operation of the drive motor at a verylargely steady-state rotation speed made possible by the presentinvention, simultaneously with the various rotation speed/torque pairsof the working apparatus which are usable, result both in a decrease infuel consumption and in reduced pollutants.

Because the drive power is the product of the rotation speed and torqueof the components transferring the drive power, the same drive power canbe achieved using different rotation speed/torque pairs that yield thesame product.

With the use of a direct-drive mode as a gearing stage in which acontinuous physical connection is established between the principalinput drive and principal output drive, the present Applicationfurthermore makes do with fewer components, or with fewer movingcomponents, in the shiftable transmission as compared with the existingart of the species recited previously, thereby additionally contributingto the robustness of the shiftable transmission. On the one hand, thecomponents involved are uniformly loaded as described above, which makesthem easier to design for appropriate loading and enhances theirrespective component service life. On the other hand, as describedabove, the fail-safe characteristics of the earth working machine as awhole are also enhanced, since the highly robust direct-drive mode, withvery little fault susceptibility because of the small number ofcomponents that suffice for constituting it, allow it to be operatedeven if a defect were to occur in the second gearing stage, and in factwith higher efficiency as compared with a gear pair or generally with astep-up or step-down in torque.

The designer is, however, free to select the transfer ratios not of twogearing stages but only of one. This disadvantage can nevertheless beaccepted in order to obtain the aforementioned advantages, since,especially in the preferred case of an internal combustion engineconstituting the drive motor and a milling tool that removes substratematerial, for example a milling drum or milling rotor, the high rotationspeed of the internal combustion engine as compared with the operatingrotation speed of the milling tool is reduced by at least one furthertransmission, with an increase in the transferred torque. When a furthertransmission of this kind is incorporated, what is important is less theabsolute values of the transfer ratios of the two gearing stages of theshiftable transmission, and more their relationship to one another. Theat least one further transmission can be designed, in terms of itstransfer ratio with regard to the transfer ratios of the shiftabletransmission, in such a way that the resulting overall transfer ratiosof all the transmissions arranged between the drive motor and workingapparatus assume desired values.

In order to allow the direct-drive mode constituted by the first gearingstage to be controllably activated and deactivated, according to anadvantageous refinement of the present invention provision can be madethat the first gearing stage encompasses a direct-drive clutch withwhich, depending on the engagement state of the direct-drive clutch, theinput shaft is connectable directly to the output shaft so as to rotatetogether at the same rotation speed, or is disconnectable from theoutput shaft. A slight slippage that may occur as torque is transferredvia the engaged direct-drive clutch is negligible in comparison withusual transfer ratios achieved and achievable by transmissionassemblages, and may therefore be left out of consideration whenassessing the rotation speed identity of the direct connection betweeninput shaft and output shaft established by the direct-drive clutch.

Not only the first gearing stage, but also the second gearing stage,must be capable of being controllably activated and deactivated.Selectable activation or deactivation of the second gearing stage canadvantageously be implemented in terms of design by the fact that thesecond gearing stage encompasses an intermediate shaft that is connectedto the input shaft via a first transmission sub-assemblage of thetransmission assemblage and is connected to the output shaft via asecond transmission sub-assemblage of the transmission assemblage, thesecond gearing stage furthermore encompassing an intermediate clutchwith which a torque transfer from the input shaft to the output shaftvia the intermediate shaft can be established or interrupted dependingon the engagement state of the intermediate clutch.

Torque is thus transferred from the input shaft to the intermediateshaft via the first transmission sub-assemblage, and torque istransferred from the intermediate shaft to the output shaft of theshiftable transmission with the second transmission sub-assemblage. Theintermediate clutch can in principle be arranged at any point along thetorque transfer path between the first and the second transmissionsub-assemblage. It can even be provided in the torque transfer path onthe input shaft or on the output shaft, for example in order toestablish and interrupt a torque transfer between one of the shaftsselected from the input shaft and output shaft and a torque-transferringcomponent respectively of the first and second transmissionsub-assemblage. In the interest of maximally efficient utilization ofavailable installation space, however, placement of the intermediateclutch on the intermediate shaft is preferred. The direct-drive clutchand intermediate clutch can thus be arranged at a sufficient distancefrom one another.

Simplified mounting of the intermediate shaft, as compared with anintermediate clutch arranged physically between the transmissionsub-assemblages, is achieved if the intermediate clutch is arrangedbetween a component of one of the transmission sub-assemblages and theintermediate shaft itself, since the intermediate shaft can then berotationally mounted using an unsplit shaft body. The mounting of twointermediate sub-shafts of a split intermediate shaft can thereby beavoided.

According to an advantageous refinement of the present invention, theearth working machine can comprise, in the torque transfer path from theshiftable transmission to the working apparatus (including therespective aforesaid limits of the path), a braking apparatus with whichthe working apparatus can be brought to a stop when the torque transferpath is interrupted, in order to avoid the transfer of undesired dragtorques.

Drag torques of this kind can occur, for example, at released clutches(such as the direct-drive clutch and/or the intermediate clutch) if theinput-side and output-side clutch parts are in frictional contactdespite being in a released state, in which ideally no torque isintended to be transferrable via the clutch.

Although the braking apparatus can in principle be arranged at any pointin the aforesaid torque transfer path, for optimum installation spaceutilization the braking apparatus is arranged in the shiftabletransmission. For optimum functionality, the braking apparatuspreferably interacts with a transmission part that is always rotatingtogether with the output shaft, including the output shaft itself. Inconsideration of the available installation space, interaction with acomponent on the output drive side of the intermediate clutch has provensuccessful. With an appropriate selection of the transfer ratio of thesecond transmission sub-assemblage, the working apparatus can be broughtto a stop there with less braking force than on the output shaft. Otherlocations for the interaction between braking apparatus and shiftabletransmission components should not be ruled out, however. For example, abrake disc of the braking apparatus can likewise be mounted on theoutput shaft so as to rotate together with it.

Although consideration can be given to providing the entire change inrotation speed and torque of the second gearing stage in only one of thetwo transmission sub-assemblages, this is not preferred because of thedifferent loading of components of the two transmission sub-assemblageswhich results therefrom. A more balanced load on the components involvedin the respective transmission sub-assemblages is possible in principlethanks to the fact that the first transmission sub-assemblage and thesecond transmission sub-assemblage each have a transfer ratio differingfrom 1.

A maximally uniform distribution of load to the two transmissionsub-assemblages in the context of the transfer of torque and rotationspeed via the second gearing stage is advantageously possible thanks tothe fact that the transfer ratio of the first transmissionsub-assemblage and the transfer ratio of the second transmissionsub-assemblage differ, based on the greater of the two transfer ratios,by no more than 3%, preferably by no more than 1.5%, particularlypreferably by no more than 0.75%.

Transmission sub-assemblages whose transfer ratios differ by no morethan 0.3 to 0.4% are very highly preferred. This is also possible interms of design, as will be shown in detail below.

In principle, the two transmission sub-assemblages can be implementedusing any desired transmission components that permit a transfer ofrotation speed and torque with modification of the rotation speed andtorque. For example, one or both of the transmission sub-assemblages cancomprise a belt drive or chain drive in order to transfer rotation speedand torque. Because of the comparatively small installation spacerequired for a high level of transferred power, however, it is preferredif the first transmission sub-assemblage is a first gear pair and if thesecond transmission sub-assemblage is a second gear pair.

In principle, consideration can be given to having the two transmissionsub-assemblages use identical gear pairs and thus an identical transferratio. This has the advantage that the two transmission sub-assemblagescan each be equipped with an identical set of gears, whichadvantageously minimizes component variation. If the use of identicalgear pairs is not possible, for example for reasons of availableinstallation space, it is preferred, in the interest of uniform loadingof the two transmission sub-assemblages, if the transfer ratios of thefirst and the second transmission sub-assemblage differ only slightly.In the case of transmission sub-assemblages embodied as gear pairs, thiscan advantageously be achieved in that the tooth counts of the smallergears of the first and the second gear pair differ by no more than twoteeth, and that the tooth counts of the larger gears of the first andthe second gear pair differ by no more than two teeth.

As is known per se, the functionality of the shiftable transmissionpresented by the present Application can extend beyond the mere transferof power from the drive motor to the working apparatus. For example,that shiftable transmission can be part of a distributor transmission inwhich power can be tapped off at several points. In a preferredrefinement of the present invention, provision is made for that purposethat the shiftable transmission comprises at least one, preferably morethan one, power takeoff.

When the shiftable transmission of the earth working machine accordingto the present invention comprises more than one power takeoff,assurance should advantageously be provided that all the power takeoffsare always being rotationally driven regardless of the activation of thefirst or second gearing stage. This can be achieved by the fact that allthe power takeoffs of the shiftable transmission are rotationally drivenvia the same transmission sub-assemblage.

This is advantageously the first transmission sub-assemblage, which isdriven directly via the input shaft of the shiftable transmission, sinceas a rule the input drive pinion, sitting on the input shaft of theshiftable transmission, of the first transmission sub-assemblage isalways rotating and the direct-drive clutch is arranged only after thefirst input drive pinion. All the power takeoffs of the shiftabletransmission are therefore preferably driven rotationally via the firsttransmission sub-assemblage. In this case, torque is transferred via thefirst transmission sub-assemblage even if the first gearing stage isactivated and the second gearing stage is deactivated. The powertransferred to any power takeoffs is, however, orders of magnitude lessthan the power transferred between the principal input drive andprincipal output drive, so that it is considerably less relevant to thedesign of the shiftable transmission than the power transferred to theprincipal output drive.

As is known per se and has already been stated above, the workingapparatus as a rule is not rotationally coupled directly to the outputshaft of the shiftable transmission. Instead, a further transfer ofdrive power, with a change in rotation speed and torque, occurs betweenthe shiftable transmission and the working apparatus. Preferably theoutput shaft is coupled in torque-transferring fashion to the workingapparatus with interposition of at least one further transmission.

As has proven successful in the past, preferably at least twotransmissions are provided between the principal output drive of theshiftable transmission and the working apparatus, one of which can be abelt drive and a further one of which can be a planetary gearsetassemblage. Both transfer at least drive power, and for simple powertransfer a change in the respective input- and output-side rotationspeed/torque value pairs can occur. That change, i.e. a transfer ratiodiffering from 1, is always the case with the planetary gearset. It canbe the case with the belt drive. The planetary gearset assemblage inparticular offers the possibility of accommodating large transfer ratiosin a comparatively small installation space. A belt drive, conversely,allows even long distances between the input shaft and output shaft tobe spanned, which creates greater freedom in terms of arranging thedrive components, namely the drive motor with shiftable transmission,and the working apparatus optionally having a further upstreamtransmission.

For example, the second gearing stage of the shiftable transmission canhave a torque transfer ratio from input side to output side of between0.6 and 0.9, preferably between 0.7 and 0.85, particularly preferablyapproximately 0.75. These values represent only exemplifying values,however, for a specific earth working machine, they depend on the drivemotor being used and its operating characteristics, on the desiredrotation speed range of the working apparatus, and on further transferratios implemented in the drive train between the drive motor andworking apparatus.

In the preferred case of a rotating milling tool constituting theworking apparatus, its desired working rotation speed range depends onits cutting circle diameter and its intended utilization situations. Ifits working apparatus is, as is preferred, provided replaceably on theearth working machine, the surface milling machine can use differentworking apparatuses to perform different tasks, which specificallyrequire different rotation speeds when the different working apparatuseshave substantially the same diameter or the same effective diameter.With the preferred exemplifying embodiment of a milling drum, the earthworking machine can take on various tasks, for example fine milling(removing the topmost layer of the road surface) or demolition of entirelayers of the road structure, by means of replaceable milling drumscarrying milling tools. Because the focus here is on correct functioningof the milling tool that removes the layer, what is critical for theparticular application is the cutting speed of the milling tool as anexample of a general working speed, for example the orbital speed of amilling bit tip as it circulates around the milling drum axis. Differentmilling tool cutting speeds are respectively advantageous for thedifferent applications. For fine milling, for example, a high cuttingspeed is desirable so as to generate a maximally homogeneous millingpattern, while a cutting speed that is as low as possible is desirablewhen removing entire layers, in order to reduce wear on the millingtools. Cutting speeds in the range from 3 to 8 m/s have provensuccessful in practical use for various milling tasks. The workingapparatuses used on a surface milling machine usually have substantiallyidentical cutting circle diameters, which as a rule differ by no morethan 5% in terms of the respectively largest available cutting circlediameter.

The possibility that the working circle diameter of differentreplaceable working apparatuses may differ, even though they areassociated with the same tasks or applications, should nevertheless notbe ruled out. Even in such a case, different rotation speeds arerequired at the working apparatus depending on its respective workingcircle diameter, in order to achieve identical effective speeds (orbitalspeed) of working tools of the working apparatus which define theworking circle diameter.

Taking as a basis, by way of example, a target effective speed range, inparticular a target cutting speed range, a designer who knows theworking point of the drive motor and the effective diameter of theworking apparatus, in particular the diameter of the cutting circle ofthe milling drum, can design the drive train having the shiftabletransmission in such a way that different effective speeds, which have adesired spacing (spread) from one another and both fall in the targeteffective speed range, are achievable with each of the two gearingstages.

If the aforementioned belt drive is provided between the shiftabletransmission and the working apparatus, said drive also preferably has atorque transfer ratio from the input side to the output side of 0.6to 1. The design of the belt drive is usually driven more by theinstallation space available to accommodate it than by a transfer ratioto be achieved with the belt drive.

If an internal combustion engine, in particular a diesel internalcombustion engine, is used as a drive motor, then in the preferredsteady-state range it rotates at a rotation speed in the range fromapproximately 1000 to 2500 rpm, in particular between 1300 and 2000 rpm.If the working apparatus is, as is preferred for the present invention,a milling tool (e.g. a milling drum or milling rotor) rotatable around amilling tool axis, a transfer of drive power of the internal combustionengine to the working apparatus with a reduction in rotation speed andan increase in torque is then desirable over the entire drive train ofthe working apparatus.

If a further transmission is interposed between the milling drum andbelt drive after the belt drive, that transmission, preferably embodiedas a planetary gearset assemblage, can have, for example, a torquetransfer ratio from input side to output side of more than 8, preferablymore than 13, but less than 28, preferably less than 23. The design ofthe further transmission is also subject to the boundary conditionspresented above in conjunction with the design of the shiftabletransmission. The drive train as a whole is intended to make possible,proceeding from a known optimum drive motor rotation speed and knowndimensions of the working apparatus, operating rotation speeds of theworking apparatus having a predefined spread in a rotation speed rangeascertained as discussed above.

For maintenance of the working apparatus, the drive train between theoutput shaft of the shiftable transmission and the working apparatus ispreferably embodied for temporary attachment of an auxiliary drive.Because of their easy accessibility, one of the belt pulleys of the beltdrive is preferably embodied for auxiliary drive attachment, for exampleby constitution of a corresponding coupling fixture thereon.

The earth working machine according to the present invention can be aroad miller, a surface miner, a surface stabilizer, or a recycler.

BRIEF DESCRIPTION OF THE DRAWINGS

The present invention will be explained in further detail below withreference to the appended drawings, in which:

FIG. 1 is an exemplifying view of an embodiment according to the presentinvention of an earth working machine of the present Application;

FIG. 2 is a schematic block view of a drive train of the earth workingmachine of FIG. 1 ,

FIG. 3 schematically depicts a first embodiment of the shiftabletransmission from the drive train of FIG. 2 ; and

FIG. 4 schematically depicts a second embodiment of the shiftabletransmission from the drive train of FIG. 2 .

DETAILED DESCRIPTION

In FIG. 1 , an earth working machine according to the present inventionis labeled generally with the number 10. Earth working machine 10,embodied in the present example as a large road milling machine, standson substrate U that is to be worked by it. The large road millingmachine has for this purpose front 12 and rear 14 drive units that canbe embodied as wheels or as crawler track units, which are known per seand will not be further discussed here.

Large road milling machine 10 comprises an operator's platform 16 fromwhich large road milling machine 10 can be controlled and operated.

Located in the region behind operator's platform 16, i.e. betweenoperator's platform 16 and the rear end of large road milling machine10, is an engine compartment 18 in which is arranged an internalcombustion engine 42 explained in further detail in conjunction withFIG. 2 , which furnishes drive power for propelling and operating largeroad milling machine 10.

Located in the region below operator's platform 16 and between the frontand rear drive units 12 and 14 in a longitudinal machine direction is amilling drum housing 20 in which a milling drum 22 constituting aworking apparatus, rotatable around a milling drum axis A orthogonal tothe drawing plane of FIG. 1 , i.e. the transverse direction of largeroad milling machine 10, is received.

With milling drum 22, substrate U can be removed to a removal depthpredetermined by modifying the height of machine frame 26 relative todrive units 12 and 14. Alternatively or additionally, milling drum 22can be received in machine frame 26 vertically adjustably relativethereto. The milled material removed by milling drum 22 is wetted inmilling drum housing 20 in order to decrease the environmental dustimpact of large road milling machine 10, and conveyed by dischargedevice 24, indicated in FIG. 1 as being merely in the vicinity of themachine, in front of large road milling machine 10. During operation,large road milling machine 10 therefore usually follows a truck whoseloading device it fills with bulk material while being operated toremove substrate U.

Numerous apparatuses on large road milling machine 10, for example thevertical adjustment system for machine frame 26 relative to drive units12 and 14, or the steering system for drive units 12 and 14 and thepropulsion system for drive units 12 and 14, are implemented by way ofhydraulic motors or hydraulic pumps. Milling drum 22, conversely, can bemechanically driven to move by way of internal combustion engine 42received in engine compartment 18. Internal combustion engine 42 is adrive power source both for the mechanically driven milling drum 22 andfor the hydraulically actuatable or drivable apparatuses of large roadmilling machine 10.

In FIG. 2 , the drive train of the working apparatus of earth workingmachine 10 of FIG. 1 , in the form of a highly schematic block diagram,is labeled with the reference character 40. Drive train 40 encompassesinternal combustion engine 42, for example in the form of a dieselinternal combustion engine, as a source of drive power. The use of anOtto-cycle engine as a drive motor is, however, also not excluded. Driveshaft 44 of internal combustion engine 42 is coupled, with interpositionof an elastic coupling 46 known per se, to input shaft 48 of a shiftabletransmission 50 described in detail below in conjunction with FIG. 3 .

With the use of shiftable transmission 50 it is possible to operate theinternal combustion engine at an optimum steady-state rotation speed andnevertheless to allow working apparatus 22, for example in the form of amilling drum or milling rotor, to operate at different rotation speedsand different torques. The steady-state operating rotation speed ofinternal combustion engine 42 can be selected for optimum performance,optimum emissions, and/or optimum consumption.

An output shaft 52 of shiftable transmission 50 is coupled, via a beltdrive 54 known per se, to the input side of a planetary gearsetassemblage 56 that is connected on the output side to working apparatus22 in torque-transferring fashion.

Belt drive 54 comprises at least two belt pulleys, one of which can beembodied for temporary attachment of an auxiliary drive so that workingapparatus 22 can be rotated at low speed for maintenance and/or repairpurposes.

As already noted above, what is depicted in FIG. 2 is merely a highlyschematic block diagram. Planetary gearset assemblage 56 can in fact bearranged at least in part, or in fact entirely, in the interior ofworking apparatus 22.

In the exemplifying embodiment depicted, working apparatus 22encompasses a hollow cylindrical milling drum on whose outer sidemilling bits are arranged, usually with interposition of bit holders orquick-change bit holders. The cavity radially inside the milling drumand surrounded by it offers space to at least partly accommodateplanetary gearset assemblage 56.

As a rule, what takes place in drive train 40 is that the rotation speedof drive shaft 44 of internal combustion engine 42 is stepped down andthe torque available at drive shaft 44 is stepped up. This means thatworking apparatus 22 rotates around its working apparatus axis moreslowly than drive shaft 44 does around its rotation axis, but with atorque that, ignoring unavoidable losses, is reciprocally greater.

Whereas, in the exemplifying embodiment presented, belt drive 54 stepstorque down and rotation speed up from the input side to the outputside, planetary gearset assemblage 56 steps torque up and steps speeddown. In the example shown in FIG. 2 , the planetary gearset assemblageis in fact the only one of the three rotation speed- andtorque-converting apparatuses 50, 54, 56 which steps rotation speed downand steps torque up. In the present exemplifying embodiment, forexample, the torque transfer ratio of the belt drive from the input sideto the output side can be selected to be between 0.78 and 0.8, and inthe present exemplifying embodiment the torque transfer ratio ofplanetary gearset assemblage 56 from the input side to the output sidecan be equal to approximately 20.5 to 20.7.

Input shaft 48 of shiftable transmission 50 constitutes, more precisely,a principal input drive of shiftable transmission 50. Output shaft 52,depicted in FIG. 2 , of shiftable transmission 50 likewise constitutes aprincipal output drive thereof.

A first embodiment of shiftable transmission 50 is depicted highlyschematically in FIG. 3 .

As depicted in FIG. 3 , input shaft 48 and output shaft 52 of shiftabletransmission 50 are selectably connectable to one another so as torotate together at the same speed, or disconnectable from one another,by way of a direct-drive clutch 60. The direct-drive clutch can be, forexample, a multi-disc clutch. By way of direct-drive clutch 60, inputshaft 48 can be connected to output shaft 52 of the shiftabletransmission to yield a shaft arrangement rotating together. Very highefficiency in terms of power transfer from internal combustion engine 42to working apparatus 22 is thereby achieved.

Direct-drive clutch 60 constitutes a first gearing stage fortorque-transferring connection of input shaft 48 and output shaft 52.With this first gearing stage, no transmission assemblage of any kind isinvolved in the transfer of drive power from input shaft 48 to outputshaft 52.

A second gearing stage that can likewise be implemented on shiftabletransmission 50 encompasses a transmission assemblage 64 that, uponactivation of the second gearing stage and deactivation of the firstgearing stage, transfers drive power from input shaft 48 to output shaft52 of shiftable transmission 50.

Transmission assemblage 64 encompasses a first transmissionsub-assemblage 66 located closer to the input side, and a secondtransmission sub-assemblage 68 located closer to the output side, ofshiftable transmission 50.

In the exemplifying embodiment according to the present invention thatis depicted, each of the two transmission sub-assemblages 66 and 68encompasses exactly one gear pair.

First transmission sub-assemblage 66 encompasses input drive pinion 70,which meshes with an output drive pinion 72 of the first transmissionsub-assemblage. In the example depicted, output drive pinion 72 isarranged on an intermediate shaft 74 so as to rotate together.

The tooth count of input drive pinion is assumed to be z₁ and the toothcount of output drive pinion 72 is assumed to be z₂; as shown in FIG. 3, z₁>z₂.

Intermediate shaft 74 furthermore carries input drive pinion 76 of thesecond transmission sub-assemblage, which meshes with output drivepinion 78 of the second transmission sub-assemblage on output shaft 52.

In the embodiment depicted, output drive pinion 78 is fixedly connectedto output shaft 52 of shiftable transmission 50 so as to rotate togethertherewith at the same speed. In the example depicted, input drive pinion76 of the second transmission sub-assemblage is selectably connectableto intermediate shaft 74 so as to rotate together, or disconnectabletherefrom, via an intermediate clutch 80.

Intermediate clutch 80 can again be a multi-disc clutch.

In a departure from what is depicted in FIG. 3 , intermediate shaft 74can also be embodied as a split intermediate shaft, in which case inputdrive pinion 76 can then be fixedly coupled to the output drive side ofintermediate shaft 74, and the input drive side and output drive side ofthe (now split) intermediate shaft can be connectable to one another inorder to rotate together, or disconnectable from one another, by way ofintermediate clutch 80.

Input drive pinion 76 of the second transmission sub-assemblage has atooth count z₃ that is greater than the tooth count z₄ of output drivepinion 78 of second transmission sub-assemblage 68.

The tooth counts z₁, z₂, z₃, and z₄ are selected so that the toothcounts of pinions of similar size, i.e. for example the tooth counts z₁and z₃ of the two input drive pinions 70 and 76, likewise differ fromone another by no more than two teeth, as also do the tooth counts z₂and z₄ of the similarly sized output drive pinions 72 and 78. It is thenthe case that the transfer ratios of the two transmissionsub-assemblages 66 and 68 can differ by no more than 1%, so that for anassumed overall transfer ratio j of transmission assemblage 64, it isapproximately true that each transfer ratio of transmissionsub-assemblages 66 and 68 is approximately the square root of j. Theresult is that at the meshing engagement points of the respectivetransmission sub-assemblages, i.e. between pinions 70 and 72 on the onehand and between pinions 76 and 78 on the other hand, the forces thatare transferred between the pinions are approximately the same, whichresults in homogeneous loading of the two transmission sub-assemblagesand thus uniform wear behavior for the entire transmission assemblage64.

With the arrangement of direct-drive clutch 60 and intermediate clutch80 as shown, shiftable transmission 50 can advantageously be shiftedbetween its two gearing stages under load. It is therefore possible toswitch, with no interruption in load, between direct drive and thetransfer ratio furnished by transmission assemblage 64.

For reliable stoppage of the working apparatus when both clutches(direct-drive clutch 60 and intermediate clutch 80) are released,shiftable transmission 50 preferably comprises a braking apparatus 81that, in the present embodiment, interacts with the output drive side ofintermediate clutch 80. A brake disc 81 a, on which braking force can beexerted by a brake caliper 81 b of braking apparatus 81, can be providedfor this purpose on the output drive side of intermediate clutch 80.

Merely for the sake of completeness, be it noted that shiftabletransmission 50 comprises four further power takeoffs 82, 84, 86, 88.

Hydraulic pumps 90, 92, 94, 96 are respectively coupled to powertakeoffs 82, 84, 86, 88. Shiftable transmission 50 is thus, in thepresent case, a pump distributor transmission.

In the example depicted, power takeoffs 82 and 84 are located on acommon first power takeoff shaft 85. Power takeoffs 86 and 88 arelocated on a common second power takeoff shaft 87.

First power takeoff shaft 85 is rotationally driven by a power takeoffpinion 89 that is connected to first power takeoff shaft 85 so as torotate together, and meshes permanently with input drive pinion 70 offirst transmission sub-assemblage 66.

Second power takeoff shaft 87 also has a second power takeoff pinion 91connected nonrotatably to it so as to rotate together. Said pinion isrotationally driven indirectly, with interposition of an intermediatepinion 93, by output drive pinion 72 of first transmissionsub-assemblage 66. This ensures that first and second power takeoffshafts 85 and 87 rotate in the same direction. It also ensures that,regardless of the engagement states of direct-drive clutch 60 andintermediate clutch 80, torque is always transferred both to first powertakeoff shaft 85 and to second power takeoff shaft 87.

A control apparatus of shiftable transmission 50 ensures exclusion ofcritical operating states, for example simultaneous torque transferengagement of both direct-drive clutch 60 and intermediate clutch 80, orbraking engagement of braking apparatus 81 when direct-drive clutch 60and intermediate clutch 80 are not both released. The control apparatusof shiftable transmission 50 can be implemented by a machine controlapparatus of earth working machine 10, or can be a control apparatusseparate therefrom, for example using respective microprocessors and/orstored-program control systems as known in the existing art.

In a departure from what is depicted in FIG. 3 , a part of intermediateshaft 74 which is constantly rotationally driven by the input shaft canalso be guided out of a housing of shiftable transmission 50 as anadditional or alternative power takeoff.

Only one of the two clutches (direct-drive clutch 60 and intermediateclutch 80) can be respectively activated for torque transfer, while theother must be deactivated. It is possible, however, to deactivate bothclutches 60 and 80 simultaneously, for example if torque is requiredonly at the power takeoffs but not at the principal output drive.

FIG. 4 depicts a second embodiment of a shiftable transmission accordingto the present invention. Identical and functionally identicalcomponents and component portions are labeled in FIG. 4 with the samereference characters as in FIG. 3 , but incremented by 100. The secondembodiment in FIG. 4 will be described below only to the extent that itdiffers from the first embodiment in FIG. 3 , the description of whichis otherwise also to be referred to for an explanation of the secondembodiment in FIG. 4 .

Shiftable transmission 150 of FIG. 4 comprises no power takeoffs, butonly the principal input drive through input shaft 148 and principaloutput drive via output shaft 152.

Intermediate clutch 180 in FIG. 4 is also, merely as an illustration ofdesign options, arranged not on intermediate shaft 174 but on outputshaft 152. Intermediate clutch 180 serves to selectably connect outputdrive pinion 178 of second transmission sub-assemblage 164 to outputshaft 152 or disconnect it therefrom.

Even though intermediate clutch 180 is arranged on output shaft 152 ofshiftable transmission 150, it is an intermediate clutch 180 forpurposes of the present Application because it is arranged in the torquetransfer path of transmission arrangement 168 and enables interruptionor establishment of a transfer of torque from input shaft 148 to outputshaft 152 via intermediate shaft 174.

In this second embodiment, braking apparatus 181 interacts with outputshaft 152, which supports a brake disc 181 so as to rotate together withit.

The invention claimed is:
 1. An earth working machine, comprising: adrive motor; a milling drum; a shiftable transmission connecting thedrive motor to the milling drum for transfer of a torque to rotate themilling drum, the shiftable transmission including: an input shaftcoupled to the drive motor; an output shaft coupled to the milling drum;a first gearing stage having a first torque transfer ratio, the firstgearing stage being configured to connect the input shaft to the outputshaft for transfer of torque and so as to move the input shaft and theoutput shaft rotationally together at the same rotation speed; and asecond gearing stage having a second torque transfer ratio differentfrom the first torque transfer ratio, the second gearing stage includinga transmission assemblage configured to connect the input shaft to theoutput shaft for transfer of torque and so as to move the input shaftand the output shaft rotationally together at different rotation speeds;and a plurality of power takeoffs configured such that all of theplurality of power takeoffs are always being rotationally driven whenthe input shaft is driven regardless of the activation of the first orsecond gearing stage; wherein the second gearing stage includes: anintermediate shaft; a first transmission sub-assemblage of thetransmission assemblage, the first transmission sub-assemblageconnecting the intermediate shaft to the input shaft; a secondtransmission sub-assemblage of the transmission assemblage, the secondtransmission sub-assemblage connecting the intermediate shaft to theoutput shaft; and an intermediate clutch having a first engagement statein which torque transfer from the input shaft to the output shaft isestablished via the intermediate shaft and the intermediate clutch, andthe intermediate clutch having a second engagement state in which torquetransfer from the input shaft to the output shaft via the intermediateshaft and the intermediate clutch is interrupted.
 2. The earth workingmachine of claim 1, wherein: all of the power takeoffs are rotationallydriven whenever the first transmission sub-assemblage is driven by theinput shaft.
 3. The earth working machine of claim 1, wherein: all ofthe power takeoffs are located upstream from the intermediate clutch. 4.The earth working machine of claim 1, wherein: the first gearing stageincludes a direct-drive clutch having a first engagement state in whichthe input shaft is connected to the output shaft via the direct-driveclutch so as to move the input shaft and the output shaft rotationallytogether at the same rotation speed, and the direct-drive clutch havinga second engagement state in which the input shaft is not connected tothe output shaft via the direct drive clutch.
 5. The earth workingmachine of claim 1, wherein: the first transmission sub-assemblage andthe second transmission sub-assemblage each have a transfer ratiodiffering from
 1. 6. The earth working machine of claim 5, wherein: thetransfer ratio of the first transmission sub-assemblage and the transferratio of the second transmission sub-assemblage differ based on thegreater of the two transfer ratios, by no more than 3%.
 7. The earthworking machine of claim 5, wherein: the transfer ratio of the firsttransmission sub-assemblage and the transfer ratio of the secondtransmission sub-assemblage differ based on the greater of the twotransfer ratios, by no more than 1.5%.
 8. The earth working machine ofclaim 5, wherein: the transfer ratio of the first transmissionsub-assemblage and the transfer ratio of the second transmissionsub-assemblage differ based on the greater of the two transfer ratios,by no more than 0.75%.
 9. The earth working machine of claim 1, wherein:the first transmission sub-assemblage includes a first gear pair; andthe second transmission sub-assemblage includes a second gear pair. 10.The earth working machine of claim 9, wherein: the first gear pairincludes a first smaller gear and a first larger gear; and the secondgear pair includes a second smaller gear and a second larger gear; atooth count of the first smaller gear and a tooth count of the secondsmaller gear differ by no more than two teeth; and a tooth count of thefirst larger gear and a tooth count of the second larger gear differ byno more than two teeth.
 11. The earth working machine of claim 1,further comprising: at least one further transmission connecting thedrive motor to the working apparatus for transfer of the torque.